Vibration control has been and remains an important field of study in engineering. Most commonly, the goal is the attenuation of the vibration of a primary system. Motivations include the reduction of dynamic stresses on machinery, the isolation of precision devices from shock and vibration, and the reduction of vibration-induced sound. Vibration absorbers are a well-known technique for achieving vibration attenuation in the presence of harmonic disturbances to a system. The greatest merit of vibration absorbers is that they are passive mechanisms that do not require external power or complex control algorithms. The greatest drawback to vibration absorbers is that they are generally effective only within a narrow band of frequencies and may cause large resonances in the frequency response at frequencies just above and below that narrow band. Adaptive absorbers have been designed that address this limitation by moving the effective operating band to match the input frequencies. The difficulty is in finding an adaptive design that is both simple to control and easily manufactured. The ideal absorber is one that incorporates a technology to eliminate moving parts and produce a design with a control logic that is easily implemented.
Vibration control techniques generally fall into one of three areas. These are passive, active, and adaptive-passive. Passive control techniques include the use of tuned-vibration absorbers (TVAs) and isolation mounts. As their name implies, passive techniques do not involve the addition of energy to a system, but rely on the inherent passive characteristics of the system to achieve a specified response. In its simplest form, as shown in FIG. 1, the TVA consists of a secondary mass 10 and spring 12 assembly attached to a primary mass 14 being driven by an external forcing function. Tuning the resonance of the secondary system 10/12 to the driving frequency will result in attenuation of the vibration of the primary mass 14. This effect is shown in FIG. 2, where the dashed line represents the frequency response of an undamped single degree of freedom (sdof) system to an external sinusoidal forcing input. The solid line shows the response with the addition of a vibration absorber with 0.1% of damping (c.sub.abs /m.sub.abs =2*.zeta.abs*.omega.abs). Also shown is a smaller dashed line that describes the response of the system when fitted with an absorber with 2% damping. For the data shown in FIG. 2, the natural frequency of the primary system is .omega.n=sqrt(2) and the absorber's undamped natural frequency is .omega.abs=1.0. The absorber's mass was chosen to be 10% of the primary system's mass. Many similar prior art plots show the case of the absorber 10/12 tuned to the natural frequency of the primary system 14. This is not believed to be a fair representation of the use of an absorber 10/12 in practice, as a primary system 14 with a resonance at an operating frequency is an obviously bad design. Rather, the expectation is that the primary system 14 will have been designed without a resonance at the driving frequency and that the absorber 10/12 will be used to attenuate the primary system's response below already non-resonant conditions.
Two significant observations are the addition of the second resonant peak below the absorber's tuned frequency and the effect that the addition of damping has on the resonant peaks and attenuation "notch." The increased damping of the 2% damped absorber 10/12 has the effect of smoothing out the response of the system, such that the resonant peaks are not so large, however it also reduces the depth of the notch where the attenuation of the primary system 14 is significant. The beneficial effect of the absorber 10/12 on the primary system 14 is defined as the reduction in the vibration of the primary system 14. If this reduction is defined as any response below the 0 dB line, then the region where the absorber 10/12 is effective is shown in FIG. 3.
Regarding the history of vibration absorbers, Frahm is credited as the inventor of the vibration absorber, with his 1911 patent. Ormondroyd and Den Hartog later gave a comprehensive treatment of the theory of vibration absorbers, including the effect of damping on absorber performance, in their 1928 paper (Ormondroyd, J. and Den Hartog, J. P., "Theory of the Dynamic Vibration Absorber." Transactions of the ASME, Applied Mechanics Division, APM-50-7, 1928, p. 9-22). Sun et al. provide a more recent study and examples of the application of passive TVAs in industry (Sun, J. Q., Jolly, M. R., and Norris, M. A., "Passive, Adaptive, and Active Tuned Vibration Absorbers--A Survey." Journal of Mechanical Design. Vol. 117B, June 1995, p. 234-242). The authors also describe the draw-back of passive TVAs in the limitation of their effectiveness to fixed bands of frequencies, as shown in FIG. 3. In the presence of uncertainties, which may include time-varying driving frequencies, the effectiveness of the TVAs is substantially reduced and may prove to have negative effects on the vibration of the primary structure.
The two main differences between active and passive control are the need for an external actuator and measurements for implementation of active control, while passive control needs neither. In vibration control, the active control techniques often involve driving the primary system in opposition to the external forcing function, such that the two forcing inputs cancel to produce no net motion of the primary mass. In general, active control techniques suffer from the requirement of input power equal to the disturbance signal. Additionally, active control schemes may require complicated matching of sensors and actuators and have the potential for adding instabilities to the system.
In contrasting passive and adaptive techniques, the passive techniques have the advantage of simplicity of design, reduced complexity, and guaranteed stability. Active techniques have the advantage of being able to control vibration across wider bands of operating frequencies.
Passive-adaptive control methods attempt to combine the positive aspects of the passive and active schemes into a single package. In general, active techniques are used to modify the passive characteristics of the primary system. In recent years, increasing research has been performed concerning the use of adaptable TVAs (ATVAs). Active modification of the absorber stiffness provides for a device that is on-line tunable for operation at different frequencies. The bandwidth of operation varies with the technique used for the active stiffness modification. A good example of a passive adaptive absorber is the design detailed in Franchek et al. (Franchek, M. A., Ryan, M. W., and Bernhard, R. J., "Adaptive Passive Vibration Control." Journal of Sound and Vibration, vol. 189, no. 5, 1995, p. 565-585). In this design, the stiffness of the absorber's spring is "dialed-in" by means of screwing the helical spring through a hole in a fixed plate. The spring stiffness is inversely dependent on the number of coils, so a softer spring may be achieved through screwing greater lengths of spring through the plate. A softer spring lowers the frequency of operation of the absorber, thereby allowing a tunable vibration absorber to be implemented.
Many adaptive passive absorber designs may suffer reliability limitations due to the complexity of their design and operation. There is therefore a need for an adaptive-passive absorber design that avoids the use of mechanisms that are unreliable and/or costly. The present invention is directed toward meeting this need.